Because of their simplicity of construction, the earliest known sliding surface bearing systems that were used in the first machines, were probably those used on wheeled carts, such as those found in the Egyptian pyramids. They consisted of a stationery shaft with substantially cylindrical surfaces located at each end, called "journal surfaces", that were supported by substantially axial cylindrical holes, called "bearing surfaces", that were formed in the center of two round discs or wheels.
Animal tallow was placed in the small clearance space that existed between these "journal" and "bearing" surfaces, to minimize their wear. These surfaces wore rapidly because they were made of wood, wood and stone, or soft metals, all of which have low wear resistance. Chariots also used such bearing systems.
With the advent of the steam engine and the development of railroad engines capable of pulling wheeled carts at high speeds, in the 1880's, there developed a great need to reduce the rate of wear of the "sliding cylindrical surface" bearing systems that were used by the engines and the load carrying carts. The resulting experimental discoveries of the engineer Beauchamp Tower, and his consultations with the mathematician Sir Osborne Reynolds, lead Reynolds to develop a theory of hydrodynamic lubrication that explained the development of significant static pressures, in the fluid film of lubricant that is within these fluid lubricated, cylindrical surface, sliding surface bearing systems, of sufficient magnitude to cause the bearing systems to run without significant contact between the journal and bearing surfaces. The resultant effect, was a greatly decreased, if not a total lack, of "material to material" contact pressures between the journal and bearing surfaces that eliminated substantial wear during the operation of such bearing systems, when certain design (unit loading and good surface geometry) and operating conditions (adequate lubricant quantity and viscosity and rotational speed), were met. Most of the remaining wear then occurred at very low rotational speeds incident to starting and stopping, and inability to keep lubricant in the clearance.
Refinements of this theory point to the need for, (a), increased precision with which the surfaces of the "journals" and the "bearings" approach perfect cylinders, and, (b), reduced "roughness" of these surfaces, so that these bearing systems could be operated with, (c), a sufficiently small diametral clearance between the journal and bearing surface, so that these bearing systems can operate at the smallest practical "minimum surface separation", which condition maximizes the load capacity by making the thickness of the lubricant film a minimum, and, (d), always having the axis of the bearing surface parallel with the axis of the journal surface.
This last requirement must be met with various radial loads on the shaft, and the resultant variety of bending of the shaft on which the journal surface is mounted, and over the range of the tolerances that determine the possible misalignment between the axes of the bearing and journal surfaces. This requirement can only be met in actual machines by mounting the bearing and journal surfaces so that they can move to a perfectly aligned condition. This introduces a complexity of design that can usually only be afforded in large machines, such as steam turbines and generators.
Thus, precise surface geometry and perfect alignment of the axes, permits the achievement of very high hydrodynamic film pressures and a correspondingly high load capacity, which makes possible a negligible rate of wear during operation at constant relative rotational speed. This is true for the absolute rotation of either the journal or the bearing surface, as long as one surface slides relative to the other.
The relation of hydrodynamic load capacity and the value of "minimum surface separation" and of "minimum film thickness", while all other factors that affect load capacity, such as the linear rubbing speed, diameter and axial length of the journal and bearing surfaces, clearance, alignment, viscosity of the lubricant, rate of lubricant supply to the clearance space, are kept constant, is shown in FIG. 1. This curve shape reveals that a reduction in "minimum film thickness" produces an exponential increase in load capacity. This rate of increase becomes proportionately greater, for a given incremental change in minimum film thickness, as the absolute value of minimum film thickness approaches zero as a limit.
Pressures as high as 20,000 lbs per square inch have been measured in practical bearings. Merely polishing a steel journal of a commercially produced electric motor bearing system, to be smoother and closer to purely cylindrical, caused a quadruple increase in load capacity, from 60 pounds to 240 pounds.
Because of the geometrical inaccuracies of practical journal and bearing surfaces, there is a difference between the minimum mechanical clearance 1 (FIG. 2) that can exist between the tops of the peaks of the roughness of the surfaces of the journal and bearing, and the average or effective thickness 2 of the film of lubricant that lies between valleys and peaks of the surfaces, and in which hydrodynamic pressures are generated. It is the slightly converging or wedge shape of the lubricant film that is in the valleys, that exist between the peaks of the surface roughness, as depicted in FIG. 2, that is required for the generation of hydrodynamic pressures within this film. It is the "dragging" or "pumping" of lubricant into this converging or wedge shaped clearance, that is caused by the relative tangential movement of the sliding surfaces, and adhesion of the lubricant to the journal and bearing surfaces, that causes pressure to be produced in the fluid film. The average velocity of the movement of lubricant into this converging wedge space is one half of the relative or tangential rubbing velocity between the surfaces.
The diametral clearance between the journal and bearing must become as small as 0.0015 times the journal diameter to make the convergence sufficiently small to permit development of significant static fluid pressures. The angle of convergence then is the order of less than 0.055 degrees. Achieving this low angle of convergence, when the journal surface is touching one side of the bearing surface, requires that a bearing surface for a 0.500" diameter journal surface have a diameter of 0.5*(1.0015)=0.50075", and thus a 0.0075" diametral clearance.
As the minimum mechanical clearance 1, see FIG. 2, through which the lubricant must pass (assuming the axial length of the bearing is infinitely long), as it circulates only circumferentially in the clearance space, becomes smaller, the resistance to this flow increases, and because the moving journal is now very effective in forcing lubricant into this converging space, when the convergence angle of the clearance space is very small, the static pressure in the lubricant is forced to become higher until it is able to make the lubricant flow through the narrow minimum mechanical clearance 1 as fast as it is brought into the convergent space. This balanced condition establishes the prevailing minimum film thickness, 2 of FIGS. 1 & 2.
When the minimum mechanical clearance 1 of FIG. 2, becomes as small as the roughness of the surfaces will allow without the tops of their roughness peaks actually touching, and when the surface roughness is very small, (about 0.000,016 inches from peak to valley), then the pressures developed in the converging film of lubricant become near their theoretical maximum value. This is the explanation for the different values and shapes of the curves 14 and 16 of FIG. 1.
In all of this discussion so far, the assumption has been made that the axis of the cylindrical journal surface, and the axis of the cylindrical bearing surface, are perfectly parallel, in all planes, and at all times, as depicted in FIG. 3. The probability that this condition prevails for just one of the bearings of a practical machine, such as for a bearing system of an electric motor or a grinding spindle, is very low, because of the range of the many tolerances that determine the principal angles between these axes and because of the bending of the shaft due to radial loads.
Now this parallelism is an unrealistically assumed condition for the "as assembled" machine, before any radial loads are applied to the shaft or rotor of the machine. Once a radial load is applied to a rotor shaft extension, it causes the journal that is nearest to the radial load, to move radially from its unloaded position toward its bearing surface at the outer end of the bearing, and the journal to become misaligned within the bearing, as in FIG. 4. All of this explanation shows that if a typical cylindrical journal surface is 0.5000" in diameter, and rotates within a cylindrical bearing surface that is 0.5008" in diameter and 0.650" in axial length, and an overhanging radial load (caused by belt tension and pulley weight) holds the journal against one side of its bearing surface, then the journal would only touch the bearing surface at a circle of contact at its outer end, instead of over its full axial length, as perfect parallelism assumes.
These bearing and journal size dimensions are typical of those of a bearing system that is produced in quantities of many millions per year for electric motors that drive outdoor condenser cooling fans for home air conditioning systems. The bending of the shaft that is caused by the above radial load thus also adds to the lack of parallelism between the bearing and journal axes.
Now the peak hydrodynamic pressures that are developed within the fluid film in such a bearing system when it satisfies the perfectly parallel requirement, only exist at the axial center of the bearing, and decrease to zero at the outer ends of the bearing, as depicted in FIG. 3, which shows a symmetrical parabolic distribution of the hydrodynamic pressures in an axial plane through the bearing axes. The total load supporting capacity of the bearing system is a maximum when this symmetrical axial pressure distribution prevails.
When the journal and bearing axes are misaligned by the maximum amount that can prevail in a given bearing system, the axial distribution of pressure becomes asymmetrical as depicted in FIG. 4. The peak pressure is now lower and the total supporting capacity of the bearing system is now much lower, as much as 60% lower than for the parallel condition.
The angle through which the circumferential distribution of pressure exists can vary greatly, depending on how fast lubricant is supplied to the converging clearance space. The maximum possible span is 180 degrees, but practically it is much less, particularly if the lubricant is supplied at a rate that is less than the converging film can utilize, (which is called "the classical feed rate").
The capillary wicking systems that are used to feed fluid lubricant to the journals of the example bearing system above, cause this angular span to be nearer 90 degrees for a new bearing, and much less as lubricant is lost from the system. The zero pressure boundaries of the pressurized film of lubricant define an elliptical shape when the area is projected onto the bearing surface.
Now the cylindrical surfaces (and their axes), of the journal and bearing are rarely, in practice, parallel which is a perfect condition. The tolerances of the dimensions of machine components, that can be economically held for mass produced machines and bearing systems, and the axial lengths of the bearing surfaces, determine the angular degree of lack of parallelism or misalignment that prevails between the journal and bearing surfaces of all practical cylindrical surface bearing systems.
In fact, the average resultant amount of misalignment cause the cylindrical journals to touch the ends of the cylindrical journals for a significant (15%) percentage of high production assemblies, such as for electric motors, and bind against them with such high forces that the frictional torque required to turn the journals in the bearings is excessive, and is large enough to cause low average bearing system life. This is true because the diametral clearances that are needed to produce the necessary low levels of quietness and high levels of hydrodynamic fluid pressure, must be so small, i.e., about 0.0015.times.the journal diameter.
The bearing surfaces have traditionally been made of a layer of the metal Tin, at least 0.020" thick, that is bonded to a backing layer of steel, about 0.060" thick. The bi-metal material is then rolled into a cylinder with the steel on the outer diameter. The outer diameter is then ground to a medium tolerance so that the cylinder can be press fitted into a machined hole in a housing. The layer of Tin at the bore is then machined to a precision diameter that is concentric to the mounting surface on the outer diameter of the housing.
One of the many virtues that is provided by the Tin bearing surface is that when the journal is highly misaligned after machine assembly, the harder steel journal surface can rapidly wear a way some of the Tin at the ends of the bearing which relieves this contact pressure, if the motor or machine has sufficient operating torque to cause rotation. In the case of inherently low torque motors, such as those used to drive direct mounted fans, the cylindrical bearing surface has traditionally been supported by a spherical outer surface that is held into a spherical socket by an axial spring, so that the bearing surface can be aligned to the shaft journal if sufficient tilting torque, to promote improved alignment, is available. Sufficient tilting torque is usually not available and the bearings actually operate in a misaligned condition.
These latter bearing systems have been called "self aligning", which is a misnomer, because all practical designs require more tilting torque to make them align than is available in the application. For instance, the bearing systems of many motors that drive automotive heating and cooling system blower wheels run misaligned throughout their entire useful life, which is often too short, because these misaligned bearings quickly loose their supply of lubricant and become noisy. This was the case with many automotive blower motors of a particular design and manufacture that had an unacceptably high field failure rate.
FIG. 5 shows that when the journal and bearing surfaces are made to be a portion of a sphere that the initial assembled relationship and also any change in the relative tilting angular relationship (such as the shaft and its affixed spherical journal moving from position 4 to position 6 as indicated by arrow 5) between the spherical journal and the adjacent spherical bearing surface cannot cause the equivalent of the misalignment that cylindrical surface bearing systems can have. The clearance between the journal and bearing versus axial location in the bearing is not altered by initial assembly tolerances or by bending of the shaft. This means that such a bearing system always has initially and always retains the inherent ability to develop its maximum hydrodynamic load capacity.
In spherical surface bearing systems, the clearance between the journal and bearing surfaces, where load causes the journal surface to be closest to the bearing surface, is not perfectly uniform initially. But calculations show that a very small amount, (about 70% of the nominal initial radial clearance), of wearing away of the bearing surface is sufficient to make this clearance uniform for any direction of loading, whether pure radial, pure axial, or a mix of these. This preferential wear of the bearing surface may be obtained by having the hardness of the journal surface greater than that of the initially hard bearing surface.
The non-uniform initial clearance does not prevent the bearing system from having appreciable load capacity, especially for radial loads, but the slight wear that makes the clearance uniform increases the load capacity, as is depicted in FIG. 5. Wear is needed more for axial loads because there is no convergence in the clearance initially, as is true for radial loads.
A test on an axially loaded spherical journal, as reported by M. C. Shaw and C. D. Strang, in the Journal Of Applied Mechanics of the American Society Of Mechanical Engineers for June 1948 and March 1949 show a surprisingly high axial load capacity is readily obtained even with a low viscosity lubricant and moderate revolutions per minute. Wear causes a relatively reduced clearance at the periphery of the bearing area that inhibits flow of lubricant out of the clearance, that enhances load capacity.